System and method for hydraulic transformer clutches

ABSTRACT

A hydraulic transformer clutch employs radial hydraulic piston assemblies with integrated electrohydraulic actuation. The hydraulic transformer clutch includes: an output shaft, an output disc affixed to the output shaft for rotation therewith, an input shaft, a rotatable housing affixed to one of the input shaft or the output shaft for rotation therewith, a plurality of hydraulic cylinders, and a plurality of working pistons. The hydraulic cylinders are operatively connected to the rotatable housing, and are spaced about the rotatable housing. Each working piston is slidably mounted within a corresponding hydraulic cylinder of the plurality of hydraulic cylinders, and is positioned to be selectively pushed, when actuated, to create a rigid connection between the input shaft and the output shaft. One or more actuator pistons are pushed by an electromagnet and create pressure that is distributed on working piston surfaces and generates active torque.

TECHNICAL FIELD

The present disclosure relates to hydraulic transformer couplingcomponents for automobiles and other wheeled vehicles.

BACKGROUND Description of the Related Art

Relatively simple hydraulic systems have been used for thousands ofyears and throughout the history of civilization, such as for irrigationand the provision of mechanical power using, for example, water wheels.In modern times, hydraulic systems have become increasinglysophisticated, and are used in a wide variety of industries for a widevariety of purposes. In general, hydraulic systems use liquids, andparticularly pressurized liquids, to generate, control, and transmitmechanical power.

A clutch is a mechanical device which engages and disengages powertransmission from a driving shaft to a driven shaft. In order totransfer torque from an internal combustion engine to a vehicletransmission, a clutch is required to engage and interrupt torquetransmission when necessary based on the operating conditions of thevehicle. Dry clutches in one or more configurations are often used toincrease the transfer surface. Complexity, torque ripple and actuationfrequently associated with wear due to dry friction are a few reasonsfor the continuing development of clutches.

Motorcycles typically employ a wet clutch, with the clutch positioned inthe same oil as the transmission. These clutches usually comprise astack of alternating plain steel and friction plates. Some platesinclude lugs on their inner surfaces that lock them to the enginecrankshaft. Wet clutches have a large engagement threshold, assuring asmooth engagement, and have higher durability and lower noise.

A dual clutch transmission (DCT) (sometimes referred to as a twin-clutchtransmission) is a type of automatic transmission or automatedautomotive transmission. A dual clutch transmission provides automaticcontrol, avoiding the use of an energy inefficient torque converter oftypical automatic transmissions. A dual clutch transmission uses a pairof clutches, one clutch engaging odd-numbered gear sets and the otherclutch engaging even-numbered gear sets. A dual clutch transmission canfundamentally be described as two separate manual transmissions (withtheir respective clutches) contained within one housing, and working asa single unit. A dual clutch transmission is usually operated in a fullyautomatic mode, providing a smooth and fast engagement of the activegear sets, well beyond the shifting capabilities of an operator. A dualclutch transmission also allows the vehicle operator to manually shiftgears in semiautomatic mode. However, the structure of the transmissionis complex, and thus expensive.

Dual clutch transmissions use two fundamentally different types ofclutches: either two wet multi-plate clutches, bathed in oil (forcooling), or two dry single-plate clutches. The wet clutch design isgenerally used for higher torque engines, whereas the dry clutch designis generally more suitable for smaller vehicles with lower torqueoutputs. However, while the dry clutch variants may be limited in torquecompared to their wet clutch counterparts, the dry clutch versions offeran increase in fuel efficiency, due to the lack of pumping losses of thetransmission fluid in the clutch housing. Wet clutches generally have awide engagement threshold and are easy to use, assuring a smoothengagement with a higher durability and lower noise than dry clutchvariations.

Currently, three variations of clutch installation are generally used. Afirst variation uses a concentric arrangement, where both clutches sharethe same plane when viewed perpendicularly from the transmission inputshaft, along the same centerline as the engine crankshaft. In order toactuate them separately, one of the clutches must be larger in diameterthan the other. As a result, different forces and control requirementsapply for the two clutches. A second variation uses a side-by-sidearrangement of two identically sized clutches around the rotational axisof the crank shaft, which increases space requirements. A thirdvariation also uses two separate but identically-sized clutches arrangedside-by-side when viewed head-on (along the length of the input shaftand crankshaft centerline), and also share the same plane when viewedperpendicularly. This clutch arrangement (unlike the other twovariations) requires an additional gear set to assure the samerotational direction of the output shaft.

The current operation of dual clutch transmissions has a number ofdisadvantages, including different drag torques in subsequent gearratios, manufacturing cost, and the amount of drag torque. Based onmanufacturing limitations, current solutions are not scalable. In orderto obtain a different transmission torque, the clutch must be resized.

Accordingly, technological improvements of the present disclosureinvolve overcoming the above shortcomings of the prior art, includingovercoming limitations related to configuration/assembly, cost, reduceddrag torque and differentiating drag torque for the differenttransmission gears involved.

BRIEF SUMMARY

A hydraulic transformer clutch employs radial hydraulic pistonassemblies with integrated electrohydraulic actuation. The hydraulictransformer clutch includes: an output shaft, an output disc affixed tothe output shaft for rotation therewith, an input shaft, a rotatablehousing affixed to one of the input shaft or the output shaft forrotation therewith, a plurality of hydraulic cylinders, and a pluralityof pistons. The hydraulic cylinders are operatively connected to therotatable housing, and are spaced about the rotatable housing. Eachpiston is slidably mounted within a corresponding hydraulic cylinder ofthe plurality of hydraulic cylinders, and is positioned to beselectively pushed, when actuated, to create a rigid connection betweenthe input shaft and the output shaft. One or more actuator pistons arepushed by an electromagnet and create pressure that is distributed onpiston surfaces and generates active torque.

The hydraulic transformer clutch may further comprise a plurality ofrotating engagement elements, each rotating engagement element of theplurality of rotating engagement elements associated with a piston ofthe plurality of pistons, wherein the plurality of rotating engagementelements engage the output disc when actuated. The hydraulic transformerclutch may further comprise a hydraulic system operatively associatedwith the hydraulic cylinders and pistons, wherein the hydraulic systemenables actuation and de-actuation of the pistons in the hydrauliccylinders, wherein actuation of the pistons in the hydraulic cylinderscouples the input shaft to the output shaft and de-actuation of thepistons in the hydraulic cylinders decouples the input shaft from theoutput shaft.

Each hydraulic cylinder of the plurality of hydraulic cylinders may bepositioned at a substantially constant offset angle relative to radialdirections of the input shaft and the output shaft. The hydraulictransformer clutch may further comprise a hydraulic variabledisplacement pump that is operatively associated with each hydrauliccylinder and piston. The hydraulic transformer clutch may furthercomprise directional control valves that use hydraulic fluid toselectively urge each piston to be pushed towards the output disc whenactuated.

The hydraulic transformer clutch may further comprise a hydraulicaccumulator that is operatively associated with the hydraulic variabledisplacement pump and directional control valves, wherein the hydraulicaccumulator reduces oscillations in the hydraulic transformer clutchduring actuation. The hydraulic transformer clutch may further comprisea pressure relief valve that protects against pressure overloads. Thehydraulic transformer clutch may be incorporated into an automotivetransmission.

A hydraulic transformer clutch system employs axial hydraulic pistonassemblies with integrated electrohydraulic actuation. The hydraulictransformer clutch system includes: an output shaft, an output discaffixed to the output shaft for rotation therewith, an input shaft, arotatable housing affixed to the input shaft for rotation therewith, aplurality of hydraulic cylinders, a plurality of working pistons, and atleast one actuation piston. The plurality of hydraulic cylinders areoperatively connected to the rotatable housing, positioned in an axialconfiguration, and spaced about the rotatable housing. Each workingpiston of the plurality of working pistons is slidably mounted within acorresponding hydraulic cylinder of the plurality of hydrauliccylinders. Additionally, each working piston of the plurality of workingpistons is positioned to be selectively pushed, when actuated, towardsthe output disc and to create a rigid connection between the input shaftand the output shaft. The at least one actuation piston is urged by asolenoid plate that includes an electromagnet solenoid. When an electriccurrent passes the electromagnet solenoid, a generated electromagneticforce pushes the solenoid plate towards a coupling housing, and the atleast one actuation piston generates, via displaced volume, a pressureincrease inside a working space associated with the plurality of workingpistons.

The hydraulic transformer clutch system may further comprise: rollerengagement elements located outwardly of the pistons, wherein thehydraulic fluid urges the roller engagement elements into engagementwith the input ring, the engagement of the roller engagement elementswith the input ring causing the output body to move in unison with theinput ring. The roller engagement elements may be cylindrical. Theoutput body arms may be angularly disposed with respect to one anotherat substantially equal angles. At least two hydraulic cylinders andassociated pistons of the plurality of hydraulic cylinders andassociated pistons may be disposed at a radially outward end of each ofthe output body arms.

The hydraulic transformer clutch system may further comprise: a gearring extending radially outwardly from the input ring for selectiveengagement with an electric motor/generator. The hydraulic transformerclutch system may further comprise: a hydraulic system operativelyassociated with the hydraulic cylinders and pistons, wherein thehydraulic system enables actuation and de-actuation of the pistons inthe hydraulic cylinders, wherein the actuation of the pistons in thehydraulic cylinders couples the input shaft to the output body and thede-actuation of the pistons in the hydraulic cylinders decouples theinput shaft from the output body. The hydraulic transformer clutchsystem may further comprise: a hydraulic variable displacement pump thatis operatively associated with each hydraulic cylinder and associatedpiston.

The hydraulic transformer clutch system may further comprise directionalcontrol valves that use the hydraulic fluid to selectively urge eachpiston to be pushed towards the output body when actuated. The hydraulictransformer clutch system may further comprise a hydraulic accumulatorthat is operatively associated with the hydraulic variable displacementpump and directional control valves, wherein the hydraulic accumulatorprovides intermediate control of the hydraulic transformer clutch systemby damping oscillations in the hydraulic transformer clutch systemduring actuation. The hydraulic transformer clutch system may beincorporated into an automotive transmission.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

In the drawings, identical reference numbers identify similar elementsor acts. The sizes and relative positions of elements in the drawingsare not necessarily drawn to scale. For example, the shapes of variouselements and angles are not necessarily drawn to scale, and some ofthese elements may be arbitrarily enlarged and positioned to improvedrawing legibility. Further, the particular shapes of the elements asdrawn are not necessarily intended to convey any information regardingthe actual shape of the particular elements, and may have been selectedsolely for ease of recognition in the drawings.

FIG. 1A illustrates a clutch with radial actuation and exterior feeding,according to at least one illustrated embodiment.

FIG. 1B illustrates a clutch with radial actuation and interior feeding,according to at least one illustrated embodiment.

FIG. 1C illustrates a clutch with axial actuation, according to at leastone illustrated embodiment.

FIG. 1D illustrates an integrated actuator and coupling device,according to at least one illustrated embodiment.

FIG. 1E illustrates a schematic diagram of an automotive transmissionincluding a brake energy recovery system, according to at least oneillustrated embodiment.

FIG. 2 illustrates a schematic diagram of an automotive transmissionimplemented with a planetary gear set including a brake energy recoverysystem, according to at least one illustrated embodiment.

FIG. 3 illustrates a schematic diagram of a hydraulic system feeding anautomotive hydraulic clutch, shown in cross-section, for use in anautomotive gearbox, according to at least one illustrated embodiment.

FIG. 4A illustrates a cross-sectional view of an automotive clutch foruse in an automotive gearbox, according to at least one illustratedembodiment.

FIG. 4B illustrates a perspective view of an automotive clutch for usein an automotive gearbox, according to at least one illustratedembodiment.

FIG. 4C illustrates a perspective view of an automotive clutch for usein an automotive gearbox, according to at least one illustratedembodiment.

FIG. 5 illustrates a schematic diagram of an automotive gearbox,according to at least one illustrated embodiment.

FIG. 6 illustrates a perspective view of a three-dimensional model of anautomotive gearbox, according to at least one illustrated embodiment.

FIG. 7A illustrates the size of a standard automotive gearbox, to becompared with the size of an automotive gearbox as described herein andillustrated in FIG. 7B, according to at least one illustratedembodiment.

FIG. 7B illustrates the size of an automotive gearbox as describedherein, to be compared with the size of the standard automotive gearboxillustrated in FIG. 7A, according to at least one illustratedembodiment.

FIG. 8 illustrates the results of a computer analysis of drag torquesresulting from gear rotation in an automotive gearbox as describedherein, according to at least one illustrated embodiment.

FIG. 9A illustrates a schematic view of the radial hydraulicpiston-actuated torque transfer device, according to at least oneillustrated embodiment.

FIG. 9B illustrates a sectional view along the plane A-A of FIG. 9A,according to at least one illustrated embodiment.

FIG. 10A illustrates a side sectional view of the dual hydraulicactuated piston clutch assembly with hydraulic seals, according to atleast one illustrated embodiment.

FIG. 10B illustrates a perspective sectional view of the dual hydraulicactuated piston clutch assembly with hydraulic seals, according to atleast one illustrated embodiment.

FIG. 11 illustrates a side sectional view of the dual hydraulic actuatedpiston clutch assembly with labyrinth seals, according to at least oneillustrated embodiment.

FIG. 12 illustrates a front sectional view of the dual hydraulicactuated piston clutch assembly with hydraulic seals, according to atleast one illustrated embodiment.

FIG. 13A illustrates a front elevational view of the dual hydraulicactuated piston clutch, according to at least one illustratedembodiment.

FIG. 13B illustrates a perspective sectional view of the dual hydraulicactuated piston clutch, according to at least one illustratedembodiment.

FIG. 14 illustrates a hydraulic actuation circuit of the two clutches,according to at least one illustrated embodiment.

FIG. 15 illustrates a schematic view of the torque transfer deviceassociated with a vehicle transmission implemented as a dual clutchtransmission, according to at least one illustrated embodiment.

FIG. 16 illustrates an exploded view of the torque transfer device shownin FIG. 9A, according to at least one illustrated embodiment.

FIG. 17 is a sectional view of the torque transfer device, according toat least one illustrated embodiment.

FIG. 18 is a sectional view showing the power/actuation hydraulic fluidpath in the torque transfer device, according to at least oneillustrated embodiment.

FIG. 19 is a cross-sectional view of the power/actuation hydraulic fluidpath in the torque transfer device, according to at least oneillustrated embodiment.

FIG. 20 is a sectional view of a dual clutch configuration of the torquetransfer device, according to at least one illustrated embodiment.

FIG. 21 is a perspective sectional view of a hydraulic piston torquetransfer device for large torque applications with multiple layers ofpistons, according to at least one illustrated embodiment.

FIG. 22 is a perspective view of a hydraulic piston torque transferdevice for large torque applications with multiple layers of pistons,according to at least one illustrated embodiment.

FIG. 23 illustrates an exploded perspective view of the multiple pistontorque transfer device, according to at least one illustratedembodiment.

FIG. 24A illustrates a front view of the multiple piston torque transferdevice shown in FIG. 23, according to at least one illustratedembodiment.

FIG. 24B illustrates a side view of the multiple piston torque transferdevice shown in FIG. 23, according to at least one illustratedembodiment.

FIG. 25 illustrates a perspective view of various components of anautomobile including a brake energy recovery system, according to atleast one illustrated embodiment.

FIG. 26 illustrates a perspective view of some of the componentsillustrated in FIG. 25, according to at least one illustratedembodiment.

FIG. 27 illustrates schematic and cross-sectional views of components ofa brake energy recovery system, according to at least one illustratedembodiment.

FIG. 28 illustrates a brake energy recovery system integrated with othercomponents of a front-wheel drive system of an automobile via a set ofgears, according to at least one illustrated embodiment.

FIG. 29 illustrates a brake energy recovery system integrated with othercomponents of a front-wheel drive system of an automobile via a shaft,according to at least one illustrated embodiment.

FIG. 30 illustrates a dual-sided brake energy recovery system integratedwith other components of a front-wheel drive system of an automobile,according to at least one illustrated embodiment.

FIG. 31A illustrates a rear-wheel drive system of an automobile,according to at least one illustrated embodiment.

FIG. 31B illustrates a brake energy recovery system in place of a driveshaft of the rear-wheel drive system illustrated in FIG. 31A, accordingto at least one illustrated embodiment.

FIG. 32A illustrates an all-wheel drive system of an automobile,according to at least one illustrated embodiment.

FIG. 32B illustrates a brake energy recovery system in place of a driveshaft of the all-wheel drive system illustrated in FIG. 32A, accordingto at least one illustrated embodiment.

FIG. 32C illustrates a Rear Wheel Drive system coupled Double ActingBrake Energy Recovery System—Structure with integrated direct mechanicaltransmission.

FIG. 32D illustrates a Rear Wheel Drive system coupled Double ActingBrake Energy Recovery System—Structure with integrated power splittransmission.

FIG. 32E illustrates an All Wheel Drive system coupled Double ActingBrake Energy Recovery System with a basic structure.

FIG. 32F illustrates an All Wheel Drive system coupled Double ActingBrake Energy Recovery System.

FIG. 32G illustrates an All Wheel Drive system coupled Double ActingBrake Energy Recovery System with a structure having an integratedhydrostatic transmission.

FIG. 32H illustrates an All Wheel Drive system coupled Double ActingBrake Energy Recovery System with a structure having an integrateddirect mechanical transmission.

FIG. 32I illustrates an All Wheel Drive system coupled Double ActingBrake Energy Recovery System with a structure having an integrated powersplit transmission and conventional gear sets.

FIG. 32J illustrates an All Wheel Drive system coupled to Double ActingBrake Energy Recovery System with a structure having an integrated powersplit transmission and planetary gear sets

FIG. 33 illustrates a variety of operating conditions for a brake energyrecovery system coupled to a drive shaft of a drive system of anautomobile, according to at least one illustrated embodiment.

FIG. 34A illustrates a drive system of an automobile, according to atleast one illustrated embodiment.

FIG. 34B illustrates two brake energy recovery systems coupled to adrive shaft of the drive system illustrated in FIG. 34A, according to atleast one illustrated embodiment.

FIG. 34C illustrates six brake energy recovery systems coupled to adrive shaft of the drive system illustrated in FIG. 34A, according to atleast one illustrated embodiment. FIG. 35A illustrates a drive system ofan automobile, according to at least one illustrated embodiment.

FIG. 35B illustrates brake energy recovery systems coupled to each ofthe axles of the drive system illustrated in FIG. 35A, according to atleast one illustrated embodiment.

FIG. 36 illustrates electrical connections of a brake energy recoverysystem, according to at least one illustrated embodiment.

FIG. 37A illustrates a brake energy recovery system, including a thermalenergy recovery subsystem, in place of an axle of a drive system of anautomobile, according to at least one illustrated embodiment.

FIG. 37B illustrates a brake energy recovery system, including a thermalenergy recovery subsystem, in place of a drive shaft of an all-wheeldrive system, according to at least one illustrated embodiment.

FIG. 37C illustrates a brake energy recovery system, including a thermalenergy recovery subsystem, in place of a drive shaft of a rear-wheeldrive system, according to at least one illustrated embodiment.

FIG. 38 illustrates a logic flow diagram for operation of a brake energyrecovery system, according to at least one illustrated embodiment.

FIG. 39 illustrates a block diagram of operation of a brake energyrecovery system, according to at least one illustrated embodiment.

FIG. 40 illustrates a block diagram of a control unit for a brake energyrecovery system, according to at least one illustrated embodiment.

FIG. 41 illustrates a block diagram of operation of a brake energyrecovery system, according to at least one illustrated embodiment.

FIG. 42 illustrates analysis of efficiency improvements provided by abrake energy recovery system, according to at least one illustratedembodiment.

DETAILED DESCRIPTION

In the following description, certain specific details are set forth inorder to provide a thorough understanding of various disclosedembodiments. However, one skilled in the relevant art will recognizethat embodiments may be practiced without one or more of these specificdetails, or with other methods, components, materials, and the like. Inother instances, well-known structures associated with the technologyhave not been shown or described in detail to avoid unnecessarilyobscuring descriptions of the embodiments.

Unless the context requires otherwise, throughout the specification andclaims that follow, the word “comprising” is synonymous with“including,” and is inclusive or open-ended (i.e., does not excludeadditional, un-recited elements or method acts).

Reference throughout this specification to “one embodiment” or “anembodiment” means that a particular feature, structure or characteristicdescribed in connection with the embodiment is included in at least oneembodiment. Thus, the appearances of the phrases “in one embodiment” or“in an embodiment” in various places throughout this specification arenot necessarily all referring to the same embodiment. Furthermore, theparticular features, structures, or characteristics may be combined inany suitable manner in one or more embodiments.

As used in this specification and the appended claims, the singularforms “a,” “an,” and “the” include plural referents unless the contextclearly dictates otherwise. It should also be noted that the term “or”is generally employed in its broadest sense, that is, as meaning“and/or” unless the context clearly dictates otherwise.

The headings and Abstract of the Disclosure provided herein are forconvenience only and do not limit the scope or meaning of theembodiments.

In some embodiment of the Hydraulic Transformer Clutch system withintegrated electrohydraulic actuation, one or more actuator pistons arepushed by an electromagnet and create pressure that is distributed onworking piston surfaces, which generates active force that in turngenerates active torque. Several different embodiments of the HydraulicTransformer Clutch system are described below.

FIG. 1A illustrates a clutch with radial actuation and exterior feeding10, according to at least one illustrated embodiment. The clutch withradial actuation and exterior feeding 10 includes a larger inlet driveand housing 12 and a smaller outlet drive 14. In this embodiment, theinlet drive and housing 12 includes an outer rim 16 with a plurality ofinwardly facing hydraulic pistons 18 mounted thereto. In thisembodiment, the outlet drive 14 includes an inner rim 20 with an outersurface. The hydraulic pistons 18 each include inwardly facingengagement elements 22 (which can be wheels) that contact the outersurface of the inner rim 20 of the smaller outlet drive 14. Thehydraulic pistons 18 are pushed by the electromagnet 24 and createpressure on working piston surfaces.

One of the technological improvements provided by this embodiment is thetorque density that is achieved by this configuration. Potentialdrawbacks that are associated with this configuration include the shapeof the pistons and that due to the pressure created by the undesirableinertial force of the pistons, the solenoid will be approximately twentypercent larger.

FIG. 1B illustrates a clutch with radial actuation and interior feeding30, according to at least one illustrated embodiment. The clutch withradial actuation and interior feeding 30 includes a smaller inlet drive32 and a larger outlet drive and housing 34. In this embodiment, theinlet drive 32 includes an outer rim 36 with an inner surface. In thisembodiment, the outlet drive and housing 34 includes an inner rim 40with an outer surface. The inner rim 40 of the outlet drive 34 includesa plurality of outwardly facing hydraulic pistons 38 mounted thereto.The hydraulic pistons 38 each include outwardly facing engagementelements 42, (which can be wheels) that contact with inner surface ofthe outer rim 36 of the smaller inlet drive 32. The hydraulic pistons 38are pushed by the electromagnet 44 and create pressure on working pistonsurfaces.

One of the technological improvements provided by this embodiment is thetorque density that is achieved by this configuration. Potentialdrawbacks that are associated with this configuration include the shapeof the pistons and that due to the pressure created by the largeundesirable inertial force of the pistons, the clutch cannot bedecoupled beyond approximately 10,000 RPMs. As a result, additionalcompensation mechanisms are employed in this configuration.

FIG. 1C illustrates a clutch with axial actuation 50, according to atleast one illustrated embodiment. The axial actuation clutch 50 includesa large inlet drive 52 and a large outlet drive 54. In this embodiment,the inlet drive 52 includes an axial rim 56 with a plurality of axiallyfacing hydraulic pistons 58 mounted thereto. In this embodiment, theoutlet drive 54 includes an angled axially rim 60 with an angledsurface. The hydraulic pistons 58 each include axially facing engagementelements 62, (which can be wheels) that contact with the angled surfaceof the angled axially rim 60 of the outlet drive 54. The hydraulicpistons 58 are pushed by the electromagnet 64 and create pressure onworking piston surfaces.

Some of the technological improvements provided by this embodiment arethe torque density that is achieved by this configuration, the shape ofthe pistons, and the fact that the pistons are not influenced byinertial forces.

FIG. 1D illustrates an integrated actuator and coupling device 70,according to at least one illustrated embodiment. In this embodiment,the integrated actuator and coupling device 70 acts as a hydraulic forceamplifier that is actuated electromagnetically. The integrated actuatorand coupling device 70 includes a large inlet shaft 72 and a largeoutput race with driven shaft 74. In this embodiment, the inlet shaft 72is associated with a coupling housing that includes an axial edge with aplurality of axially facing hydraulic working pistons 76 mountedthereto. In this embodiment, the large output race with driven shaft 74includes an angled axially rim 78 with an angled surface. The hydraulicpistons 76 each include axially facing engagement elements 80, (whichcan be wheels or other rolling elements) that contact with the angledsurface of the angled axially rim 78 of the large output race withdriven shaft 74.

Actuation pressure is generated using the hydraulic transformer concept(i.e., the Pascal principle) in which a small diameter piston isactuated with low force. This action generates pressure that isdistributed on the working piston surfaces, and consequently amplifiesthe actuation force transferred by piston and roller to the output raceand driven shaft.

The actuation piston 82 is moved by a solenoid plate 84 containing asolenoid 86. When an electric current passes through the solenoid 86,the generated electromagnetic force 87 pushes the solenoid plate 84towards the coupling housing, and the attached actuation piston 82generates, via the displaced volume, a pressure increase inside theworking space 88. The electrical energy supply of the rotating solenoid86 is done using the electrical fixed contact element 90 that is intouch with the contact race 92. The coupling fixed housing 94 issupported against the moving disc by gliding bearings 95 that compensatefor axial forces. Splashing blades 96A create an oil fog to lubricaterollers and output race. In another aspect, cooling blades 96B supportthe solenoid waste of heat dissipation. In still another aspect, aspring element 98 is enclosed into a damping cavity 99 to assure maximumpressure.

FIG. 1E illustrates an automotive transmission 100. Transmission 100includes a primary drive shaft 102, which is driven by an internalcombustion engine, electric motor, hydraulic propulsion system, or othersuitable torque or power source. The primary drive shaft 102 is rigidlycoupled to a primary drive gear 104 having primary drive gear teeth 106.Transmission 100 also includes a primary driven shaft 108, which drivesa mechanical device or system such as the wheels of an automobile. Theprimary driven shaft 108 is rigidly coupled to a primary driven gear 110having primary driven gear teeth 112.

Power is transmitted through the transmission 100 from the primary driveshaft 102 and the primary drive gear 104 to the primary driven gear 110and the primary driven shaft 108 in one or more of three independentways, depending on the operation of the transmission 100. First, thetransmission 100 includes a clutch 114 that can be engaged to directlyand rigidly couple the primary drive shaft 102 to the primary drivenshaft 108, so that the primary drive shaft 102 is mechanically locked tothe primary driven shaft 108 and so that the primary driven shaft 108turns at the same speed as the primary drive shaft 102. The clutch 114can also be disengaged to de-couple the primary drive shaft 102 from theprimary driven shaft 108, so that the primary drive shaft 102 is notmechanically locked to the primary driven shaft 108. The clutch 114 hasa structure matching or similar to that of the clutch 200 describedherein.

Second, the transmission 100 includes a first hydraulic power transfersystem 116 that is used to transfer power or torque hydraulically fromthe primary drive shaft 102 and primary drive gear 104 to the primarydriven gear 110 and primary driven shaft 108. The first hydraulic powertransfer system 116 includes a secondary drive gear 118 a havingsecondary drive gear teeth 120 a meshed with the primary drive gearteeth 106. The secondary drive gear 118 a is rigidly coupled to asecondary drive shaft 122 a, which is coupled to a hydraulic pump 124 a.Rotation of the secondary drive shaft 122 a actuates operation of thehydraulic pump 124 a, such as to increase a pressure of a hydraulicfluid or to create a hydraulic fluid pressure differential.

The first hydraulic power transfer system 116 also includes a secondarydriven gear 126 a having secondary driven gear teeth 128 a meshed withthe primary driven gear teeth 112. The secondary driven gear 126 a isrigidly coupled to a secondary driven shaft 130 a, which is coupled to ahydraulic motor 132 a. The motor 132 a is actuated to operate by theprovision of a relatively high-pressure hydraulic fluid, or a hydraulicfluid pressure differential, to drive rotation of the secondary drivenshaft 130 a and the secondary driven gear 126 a. The hydraulic pump 124a is hydraulically coupled to the hydraulic motor 132 a, so that thepressurized hydraulic fluid generated by the hydraulic pump 124 a isused to drive operation of the hydraulic motor 132 a. In someimplementations, the hydraulic pump 124 a has the same or a similarstructure as the hydraulic motor 132 a, which is a vane-type hydraulicmotor, however the hydraulic pump 124 a operates in reverse of thehydraulic motor 132 a.

Third, the transmission 100 includes a second hydraulic power transfersystem 134 that is used to transfer power or torque hydraulically fromthe primary drive shaft 102 and primary drive gear 104 to the primarydriven gear 110 and primary driven shaft 108. The second hydraulic powertransfer system 134 includes a secondary drive gear 118 b havingsecondary drive gear teeth 120 b meshed with the primary drive gearteeth 106. The secondary drive gear 118 b is rigidly coupled to asecondary drive shaft 122 b, which is coupled to a hydraulic pump 124 b.Rotation of the secondary drive shaft 122 b actuates operation of thehydraulic pump 124 b, such as to increase a pressure of a hydraulicfluid or to create a hydraulic fluid pressure differential.

The second hydraulic power transfer system 134 also includes a secondarydriven gear 126 b having secondary driven gear teeth 128 b meshed withthe primary driven gear teeth 112. The secondary driven gear 126 b isrigidly coupled to a secondary driven shaft 130 b, which is coupled to ahydraulic motor 132 b. The motor 132 b is actuated to operate by theprovision of a relatively high-pressure hydraulic fluid, or a hydraulicfluid pressure differential, to drive rotation of the secondary drivenshaft 130 b and the secondary driven gear 126 b. The hydraulic pump 124b is hydraulically coupled to the hydraulic motor 132 b, so that thepressurized hydraulic fluid generated by the hydraulic pump 124 b isused to drive operation of the hydraulic motor 132 b. In someimplementations, the hydraulic pump 124 b has the same or a similarstructure as the hydraulic motor 132 b, which is a vane-type hydraulicmotor, however the hydraulic pump 124 b operates in reverse of thehydraulic motor 132 b.

The transmission 100 also includes a high-pressure accumulator 136 and alow-pressure accumulator 138. As illustrated in FIG. 1E, thehigh-pressure accumulator 136 is hydraulically coupled by a first valve140 and a first series of hydraulic conduits 142 to an output of thehydraulic pump 124 a and to an input of the motor 132 a, and by a secondvalve 144 and a second series of hydraulic conduits 146 to an output ofthe hydraulic pump 124 b and to an input of the motor 132 b. Further,the low-pressure accumulator 138 is hydraulically coupled by a thirdseries of hydraulic conduits 148 to an input of the hydraulic pump 124 aand to an output of the motor 132 a, and by a fourth series of hydraulicconduits 150 to an input of the hydraulic pump 124 b and to an output ofthe motor 132 b. A relief valve 152 couples the high-pressureaccumulator 136 and its hydraulic conduits to the low-pressureaccumulator 138 and its hydraulic conduits to prevent excessive pressuredifferentials between the high-pressure and low-pressure accumulators.

In some implementations, an internal combustion engine driving theprimary drive shaft 102 operates continuously at its optimal or mostefficient operating parameters, independent of the power demanded at thedriven shaft 108, to improve overall system efficiency. When powersupplied by the drive shaft 102 matches the power demanded at the drivenshaft 108, the clutch 114 is engaged and the first and second hydraulicpower transfer systems 116 and 134 are disengaged. When power suppliedby the drive shaft 102 exceeds the power demanded at the driven shaft108, the hydraulic pumps 124 a and/or 124 b are operated to pumphydraulic fluid from the low-pressure accumulator 138 to thehigh-pressure accumulator 136, to store excess energy for later use.During such operations, the clutch 114 can be either engaged ordisengaged. When power supplied by the drive shaft 102 is less than thepower demanded at the driven shaft 108, the hydraulic motors 132 aand/or 132 b are actuated by high-pressure hydraulic fluid stored in thehigh-pressure accumulator 136 to power rotation of the secondary drivenshafts 130 a and/or 130 b, and to power rotation of the primary drivenshaft 108. During such operations, the clutch 114 can be either engagedor disengaged.

In some implementations, the hydraulic pumps 124 a and/or 124 b driveoperation of the hydraulic motors 132 a and/or 132 b directly, ratherthan indirectly through collection of high-pressure hydraulic fluid inthe high-pressure accumulator 136. In some implementations, thehydraulic pumps 124 a and 124 b, and/or the hydraulic motors 132 a and132 b, are of different sizes and operating capacities, so that they areoperable independently of one another based on the running conditions ofthe system supplying power to the primary drive shaft 102 and the systemdrawing power from the primary driven shaft 108, to improve overallefficiency.

In some implementations, to apply a braking or decelerating torque,rather than a driving torque, to the primary driven shaft 108, operationof the hydraulic motors 132 a and/or 132 b is inverted so the hydraulicmotors operate as hydraulic pumps. During such operations, the clutch114 is disengaged. The hydraulic pumps 132 a and/or 132 b are thenactuated by the primary driven shaft 108 and operated to pump hydraulicfluid from the low-pressure accumulator 138 to the high-pressureaccumulator 136, to store excess energy for later use.

FIG. 2 illustrates the automotive transmission 100, with somemodifications to the implementation illustrated in FIG. 1. Asillustrated in FIG. 2, the transmission 100 includes a planetary gearset 154, between the primary drive shaft 102 and the primary drive gear104. The planetary gear set 154 includes first and second planetarygears 158 a and 158 b rotatably mounted on a planetary gear carrier 156rigidly coupled to the primary drive shaft 102. The planetary gear set154 also includes a sun gear 160 rigidly coupled to the primary drivenshaft 108. The planetary gear set 154 also includes the primary drivegear 104, which functions as the planetary ring gear of the planetarygear set 154.

In some implementations, the planetary gear set performs the function ofthe clutch 114 illustrated in FIG. 1. For example, the primary drivegear 104 can be held stationary in a fixed position so that the primarydrive shaft 102 drives the primary driven shaft 108 purely mechanically,without the action of any intermediate hydraulic components. The primarydrive gear 104 can also be allowed to rotate about its own axis, so thatthe hydraulic pumps 124 a and/or 124 b, and/or the hydraulic motors 132a and/or 132 b, operate as described above with respect to FIG. 1.

FIG. 2 also illustrates that the transmission 100 includes a tertiarydrive gear 162 having teeth meshed with outer teeth of the primary drivegear 104. The tertiary drive gear 162 is rigidly coupled to a tertiarydrive shaft 164 and thereby to a hydraulic pump 166. The hydraulic pump166 is coupled to a hydraulic conduit, forming a closed loop with a flowcontrol valve 168. In some implementations, the flow control valve 168is closed to hold the primary drive gear 104 stationary in a fixedposition, and opened to allow the primary drive gear 104 to rotate aboutits own axis. In some implementations, the hydraulic pump 166 and theflow control valve 168 are operated to dampen vibrations within thetransmission 100.

FIG. 3 illustrates a hydraulically powered clutch 200 for transferringpower or torque from an input shaft 202 to an output shaft 204 in anautomotive gearbox. The clutch 200 includes a portion of the input shaft202 mounted to rotate within, and supported by, a stationary bearing206, and a portion of the output shaft 204 mounted to rotate within, andsupported by, a stationary housing 208. The input shaft 202 is rigidlycoupled to an input disc 210 by a plurality of bolts or other fasteningmembers 212. The input disc 210 is rigidly coupled to an outer rim 214,such as by welding or by a suitable adhesive. The outer rim 214 is agear having teeth 216 formed integrally in an outer surface thereof.

The output shaft 204 is integrally formed with an output disc 218 thatextends radially outward from the end portion of the output shaft 204and that is arranged parallel to the input disc 210. The input disc 210includes a recess extending into its front surface along its centrallongitudinal axis, and the output disc 218 includes a protrusionextending outward from its front surface along its central longitudinalaxis. The protrusion of the output disc 218 is seated within the recessof the input disc 210 with a bearing 230 positioned between an outersurface of the protrusion and an inner surface of the recess.

The output shaft 204 and the output disc 218 include a network ofhydraulic conduits. These hydraulic conduits extend from an outersurface of the output shaft 204 radially inward to a centerline of theoutput shaft 204 at a radial hydraulic conduit 220, along the centerlineof the output shaft 204 into the output disc 218 at a radial hydraulicconduit 222 (capped by a plug 240), and radially outward through theoutput disc 218 at a radial hydraulic conduit 224, to form a hydrauliccylinder housing a hydraulic piston 226. The hydraulic piston 226 isengaged at a first end thereof with the hydraulic fluid within thehydraulic conduits, and is coupled at a second end thereof opposite tothe first end to an engagement element 228, which can be a wheel 228, incontact with an inner surface of the outer rim 214.

The output disc 218 includes a groove adjacent to, and extendingcircumferentially around, the output shaft 204, that extends into a rearsurface of the output disc 218. The housing 208 extends into and isseated within the groove in the rear surface of the output disc 218 witha bearing 232 positioned between an outer surface of the housing 208 andan inner surface of the groove. A bearing 236 is rigidly coupled to aninner surface of the housing 208 and is engaged with an outer surface ofthe output shaft 204. The bearing 236 includes an annular groove 238that extends around the output shaft 204 and that is in hydrauliccommunication with the radial hydraulic conduit 220 within the outputshaft 204. The bearing 236 also includes a port that couples its annulargroove 238 to a hydraulic conduit 234 extending radially outward throughthe housing 208.

The radial hydraulic conduit 234 of the housing 208 is hydraulicallycoupled, such as by a hydraulic connector 242, to a hydraulic conduit244, a hydraulic flow control valve 246, an accumulator 248, which isused for storage of hydraulic energy and damping of hydraulic shocks, ahydraulic pump 250, and a hydraulic reservoir 252. In someimplementations, the valve 246 is kept closed so that the hydraulicfluid within the hydraulic conduits within the output shaft 204 and theoutput disc 218 is not highly pressurized, and so that the engagementelement 228 does not engage the inner surface of the outer rim 214. Insuch implementations, the input disc 210 rotates freely with respect tothe output disc 218, and neither power nor torque is transferred fromthe input shaft 202 to the output shaft 204.

In other implementations, the valve 246 is opened and the hydraulic pump250 is actuated to pump hydraulic fluid and generate a high-pressurehydraulic wave that travels through the valve 246 and the varioushydraulic conduits into the hydraulic cylinder within the output disc218. In other implementations, the valve 246 is opened and thehigh-pressure hydraulic fluid held within the hydraulic accumulator 248is released to generate a high-pressure hydraulic wave that travelsthrough the valve 246 and the various hydraulic conduits into thehydraulic cylinder within the output disc 218. In such implementations,when the high-pressure hydraulic wave reaches the hydraulic piston 226,it pushes the hydraulic piston 226 and the engagement element 228radially outward so that the engagement element 228 engages the innersurface of the outer rim 214. In such implementations, the input disc210 does not rotate freely with respect to the output disc 218, and isinstead rotationally locked to the output disc 218, so that power ortorque is transferred from the input shaft 202 to the output shaft 204.

FIGS. 4A and 4B illustrate a schematic side view and a perspective view,respectively, of the hydraulically powered clutch 200, with somemodifications to the implementation illustrated in FIG. 3. Asillustrated in FIGS. 4A and 4B, the clutch 200 includes the radialhydraulic conduit 224 extending radially outward from the hydraulicconduit 222 to two hydraulic transformers 254. Each hydraulictransformer 254 is used to step hydraulic pressure up from a firsthydraulic pressure within the radial hydraulic conduit 224 at a firstside of the hydraulic transformer 254 to a second hydraulic pressure ata second side of the hydraulic transformer 254 opposite to its firstside, which is in hydraulic communication with a respective hydraulicpiston 226. In some implementations, the hydraulic transformers 254 areused to step hydraulic pressure up from about 18 bar to about 100 bar.

An implementation of the hydraulically powered clutch 200 illustrated inFIGS. 4A and 4B includes two hydraulic pistons 226 and two correspondingengagement elements 228. In various implementations, the engagementelements 228 may be cylinders, wheels, spherical elements, or the like.The pistons 226 and corresponding engagement elements 228 are orientedobliquely (e.g., at an oblique angle theta) to a radial axis 260, butare rotationally symmetric with respect to the central longitudinal axis270 of the clutch 200, such that the engagement elements 228 and theforces they exert against the outer rim 214 are balanced.

FIG. 4C illustrates a schematic perspective view of the hydraulicallypowered clutch 200, with some modifications to the implementationsillustrated in the preceding figures. As illustrated in FIG. 4C, theclutch 200 includes a plurality of hydraulic pistons 226 (twenty shown)and a corresponding plurality of engagement elements 228 (twenty shown).Again, since the twenty hydraulic pistons 226 and the twentycorresponding engagement elements 228 are oriented obliquely to radialaxis 260 but radially symmetric with respect to the central longitudinalaxis of the clutch 200, they and the forces they exert against the outerrim 214 are balanced. Increasing the number of hydraulic pistons 226 andthe number of engagement elements 228 included within the clutch 200increases the overall torque and power that the clutch 200 is capable oftransferring without increasing its overall size or the complexity ofits manufacture. Thus, the clutch design is scalable for use with heavyduty vehicles such as agricultural equipment, construction equipment,and the like.

FIG. 5 illustrates an automotive gearbox 300. The automotive gearbox 300includes an input shaft 302 rigidly coupled to a first hydraulicallypowered clutch 304 and to a second hydraulically powered clutch 306. Thefirst and second hydraulically powered clutches 304 and 306 havestructures matching or similar to the clutch 200 described herein. Thefirst clutch 304 includes a first drive gear 308 that can be locked toor released from the input shaft 302 by the first clutch 304, and thesecond clutch 306 includes a second drive gear 310 that can be locked toor released from the input shaft 302 by the second clutch 306.

The first drive gear 308 has teeth meshed with the teeth of a firstdriven gear 312 rigidly coupled to an output shaft 316, and the seconddrive gear 310 has teeth meshed with the teeth of a second driven gear314 rigidly coupled to the output shaft 316. The first drive gear 308has a larger radius than the second drive gear 310, and the first drivengear 312 has a smaller radius than the second driven gear 314, so thatthe gearbox 300 has a different gear ratio when the first clutch 304 isengaged than when the second clutch 306 is engaged. Either one, but onlyone, of the first and second clutches 304 and 306 is engaged at a giventime, to prevent binding of, and damage to, the gearbox 300.

FIG. 5 also illustrates that the gearbox 300 includes a hydraulicaccumulator 318 hydraulically coupled to a hydraulic pump 320, which iscoupled to be driven by the driven shaft 316, and a hydraulic motor 322,which is coupled to drive the driven shaft 316, and which collectivelyoperate in the same ways described herein for automotive transmission100. The gearbox 300 has various technological improvements overtraditional automotive gearboxes. As one example, the gearbox 300 isrelatively short and compact. As another example, the gearbox 300 doesnot need a flywheel. As other examples, the gearbox 300 has relativelysmooth engagement, actuation, and operation. As another example, thegearbox 300 is relatively low-cost. As another example, the gearbox 300has large torque ranges. As another example, the gearbox 300 is easilyintegrated with a brake energy recovery system.

FIG. 6 illustrates a perspective view of a three-dimensional model ofthe gearbox 300, with some modifications to the implementationsillustrated in FIG. 5. As illustrated in FIG. 6, the gearbox 300includes the input shaft 302, the output shaft 316, a plurality of drivegears 330, a plurality of driven gears 332, the hydraulic pump 320, thehydraulic motor 322, and the hydraulic accumulator 318. Each of thedrive gears 330 includes a clutch having a structure matching or similarto that described herein for clutch 200.

As also illustrated in FIG. 6, the hydraulic accumulator 318 of thegearbox 300 includes a high-pressure accumulator 324 and a low-pressureaccumulator 326, with a flexible wall 328 separating the high-pressureaccumulator 324 from the low-pressure accumulator 326. The flexible wall328 deforms as the pressure of a hydraulic fluid within thehigh-pressure accumulator 324 increases and/or as the pressure of ahydraulic fluid within the low-pressure accumulator 326 decreases, tostore energy for later use, as described herein. In addition to theflexible wall 328, in some implementations, the accumulator 318 includesa pressurized gas and/or a mechanical spring to store energy for lateruse. One or both of the accumulators 324, 326 included in the hydraulicaccumulator 318 may be used to store energy recovered from a hydraulicbraking system. The accumulator 318 thus increases energy efficiency ofthe torque transfer system compared with a conventional clutch system.

FIG. 7A illustrates a traditional gearbox 400. FIG. 7B illustrates thegearbox 300, with some modifications to the implementations illustratedin the preceding figures. As illustrated in FIG. 7B, the space occupiedby the gears of the gearbox 300 is significantly shorter and smallerthan that for the traditional gearbox 400, leaving space in the gearbox300 for the hydraulic pump 320, motor 322, and/or accumulator 318without increasing the overall dimensions needed for the gearbox 300.FIG. 8 illustrates the results of computer analysis of drag torquesresulting from gear rotation in the gearbox 300. In particular, FIG. 8illustrates that the drag force decreases on the shaft carrying thedrive gears 330, which are equipped with the clutches described herein,and that the drag force increases on the shaft carrying the driven gears332, which are not equipped with the clutches described herein.

Referring now to FIG. 9A, a schematic view of the radial hydraulicpiston-actuated torque transfer device 900 is shown that displays aninput shaft 902, which is rigidly attached to a rotatable housing 904.The rotatable housing 904 contains an annular hydraulic feeding channel906 that is connected to radial feeding channels 908 of hydrauliccylinders 910.

As shown in FIG. 9B, the hydraulic cylinders 910 are positioned at aconstant offset angle relative to the radial axis. More specifically,the hydraulic cylinders 910 are canted or angularly offset relative toradial lines drawn from the coincident centers of the correspondingshafts. In some implementations, the offset angle is less than aboutthirty degrees. In more preferable implementations, the offset angle isfrom about five degrees to about twenty-five degrees.

Referring again to FIG. 9A, within each hydraulic cylinder 910 is apiston 912 gliding on a rolling engagement element 914. The rollingengagement element 914 is in contact with an output disc 916. The outputdisc 916 is rigidly fixed to an output shaft 918. The piston stroke ofeach piston 912 within each hydraulic cylinder 910 is relatively short.Notably, the rolling engagement element 914 reduces friction during theengagement process, before full coupling occurs. The force transmissionfrom the piston 912 to an inner ring (i.e., the output disc 916) isaccomplished using two contact surfaces; (1) the contact surface betweenthe piston 912 and the rolling engagement element 914, and (2) thecontact surface between the rolling engagement element 914 and the innerring (i.e., the output disc 916). This component configuration producesa more efficient overall contact on these two microscale adaptableelements.

The radial hydraulic piston-actuated torque transfer device 900 furtherincludes a fixed housing 920 that contains a hydraulic fluid coupling922 to the annular hydraulic feeding channel 906. The hydraulic feedingcoupling 922 is connected to a hydraulic circuit, which is furtherdescribed with reference to FIG. 14. The rotatable housing 904 and thefixed housing 920 are sealed by seals 924 and 926. The radial hydraulicpiston-actuated torque transfer device 900 also includes bearing 928that is positioned between the input shaft 902 and the output disc 916,which is affixed to the output shaft 918. The bearing 928 enables theinput shaft 902 and the output shaft 918 to run concentrically. Due toinertial forces, the pistons 912 are not initially in contact with theoutput disc 916. When actuated, hydraulic fluid enters the radialhydraulic piston-actuated torque transfer device 900 by way of thehydraulic fluid coupling 922. The hydraulic fluid enters the annularhydraulic feeding channel 906 that is connected to the radial feedingchannels 908 and the hydraulic cylinders 910. The hydraulic fluid thenpushes the pistons 912 and the attached rolling engagement elements 914to engage with the output disc 916, thereby transferring the torque tothe output shaft 918. Thus, mechanical torque is transferred from aninput shaft 902 to an output shaft 918 based on the friction created bythe pistons 912 that are actuated by pressurized hydraulic fluid.

Referring now to FIG. 10A, a cross-sectional view of a dual hydraulicactuated piston clutch transmission with hydraulic seals is shown.Correspondingly, FIG. 10B shows a perspective sectional view of a dualhydraulic actuated piston clutch transmission. The dual hydraulicactuated piston clutch transmission combines two of the above-describedradial hydraulic piston-actuated torque transfer devices 900. Thehydraulic fluid is controlled by the hydraulic circuit (as describedwith reference to FIG. 14). The hydraulic fluid enters the systemthrough hydraulic fluid couplings 922, and passes into fluid chambers934 and 936. The pressure generated by the hydraulic fluid within thefluid chambers 934 and 936 presses the pistons 938 and the rollingengagement elements 940 toward the output shaft assembly 942 or 944until the rolling engagement elements 940 are in contact with the outputshaft assembly 942 or 944. The contact pressure of the rollingengagement elements 940 against the output shaft assembly 942 or 944 isdirectly proportional to the pressure within the fluid chambers 934 and936. If there is sufficient pressure flow in the system, the torque frominput shaft assembly 942 or 944 is diverted to output shaft assembly946. However, if there is insufficient pressure flow, the torque frominput shaft assembly 942 or 944 acts as a free wheel. In someimplementations, as shown in FIG. 10, the system may be sealed bygaskets 948. In another aspect of some implementations, one or moreorifices 952 are included at the bottom of the system through whichleaked fluid may drain.

Referring now to FIG. 11, a side sectional view of the dual hydraulicactuated piston clutch assembly is shown with labyrinth seals 950. Insuch an implementation, the system may be sealed by using a labyrinthsealing system of known design. To prevent parasite pressure flow, anyoil leaked will drain through orifice 952. The connection to the startermotor is completed through gearing 954. This gearing 954 is shown ingreater detail in FIG. 13. The remainder of the dual hydraulic actuatedpiston clutch assembly with labyrinth seals 950 is similar to the dualhydraulic actuated piston clutch transmission with hydraulic seals. Assuch, in the dual hydraulic actuated piston clutch assembly withlabyrinth seals 950, the hydraulic fluid enters the system throughhydraulic fluid couplings 922, and passes into fluid chambers 934 and936, after which the pressure generated by the hydraulic fluid withinthe fluid chambers 934 and 936 presses the pistons 938 and the rollingengagement elements 940 towards the output shaft assembly 944 until therolling engagement elements 940 are in contact with the output shaftassembly 942 or 944. Again, if there is sufficient pressure flow in thesystem, the torque from input shaft assembly 942 or 944 is diverted tooutput shaft assembly 946.

FIG. 12 illustrates a front sectional view of the dual hydraulicactuated piston clutch assembly with hydraulic seals, according to atleast one illustrated embodiment. As shown in FIG. 12, in someimplementations, the offset angle is less than about thirty degrees. Inmore preferable implementations, the offset angle is from about fivedegrees to about twenty degrees. In the implementation shown in FIG. 12,the output disc 916 and the output shaft 918 are shown. Additionally,eighteen pistons 912 and eighteen associated rolling engagement elements914 are shown that engage with the output disc 916 when actuated bypressurized hydraulic fluid. The fixed housing 920 of the radialhydraulic piston-actuated torque transfer device 900 is also shown.Additionally, the hydraulic fluid couplings 922 are shown in FIG. 12,through which the hydraulic fluid enters the system to actuate thepistons 912 and associated rolling engagement elements 914 against theoutput disc 916.

Referring now to FIG. 13A, a front elevational view of the dualhydraulic actuated piston clutch assembly is shown. FIG. 13B shows asectional perspective view of the dual hydraulic actuated piston clutchassembly. In the implementation shown in FIGS. 13A and 13B, the outputshaft 918 and the fixed housing 920 are shown. The gearing 954 is alsoshown around the perimeter of the fixed housing 920. In the sectionalperspective view of FIG. 13B, both clutches of the dual clutch systemare shown, as well as the pistons 912 and associated rolling engagementelements 914 that engage with the output disc 916 when actuated bypressurized hydraulic fluid. Additionally, the hydraulic fluid couplings922 are shown through which the hydraulic fluid enters the system toactuate the pistons 912.

Referring now to FIG. 14, a hydraulic circuit for the dual hydraulicactuated piston clutch assembly 1400 is shown. The hydraulic circuitincludes a variable displacement pump 956 that directs hydraulic fluidfrom the tank 958 to a first directional control valve 960. The firstdirectional control valve 960 controls the flow of the hydraulic fluidthat arrives at the hydraulic fluid coupling 922 of the first radialhydraulic piston-actuated torque transfer device 900 using pipe 962. Thehydraulic fluid passes through the hydraulic fluid coupling 922 andurges the piston 938 and rolling engagement elements 940 towards theoutput shaft assembly 942 until the rolling engagement elements 940 arein contact with and rotate the output shaft assembly 942. The fixedhousing 920 of the radial hydraulic piston-actuated torque transferdevice 900 is also shown. In the implementation shown in FIG. 14, thefirst radial hydraulic piston-actuated torque transfer device 900 isintegrated into the dual hydraulic actuated piston clutch assembly 1400.

The variable displacement pump 956 of the hydraulic circuit also directshydraulic fluid from the tank 958 to a second directional control valve968. The second directional control valve 968 controls the flow of thehydraulic fluid through the pipe 970 that arrives at the hydraulic fluidcoupling 922 of the second radial hydraulic piston-actuated torquetransfer device 900. The hydraulic fluid passes through the hydraulicfluid coupling 922 and urges the piston 938 and rolling engagementelements 940 towards the output shaft assembly 944 until the rollingengagement elements 940 are in contact with and rotate the output shaftassembly 944. In the implementation shown in FIG. 14, the second radialhydraulic piston-actuated torque transfer device 900 is also integratedinto a dual hydraulic actuated piston clutch assembly.

Referring still to FIG. 14, the hydraulic circuit further includesaccumulator 976 that reduces oscillations during actuation of thepistons 912 and associated rolling engagement elements 914 using thehydraulic fluid. Additionally, the accumulator 976 provides a liquidvolume that is disposed for fast response during actuation of the radialhydraulic piston-actuated torque transfer device 900. Further, thehydraulic circuit includes a pressure relief valve 978 that protectsagainst pressure overloads in the system.

Referring now to FIG. 15, a schematic view of the radial hydraulicpiston-actuated torque transfer device 900 is shown. In thisimplementation, the radial hydraulic piston-actuated torque transferdevice 900 is associated with a dual hydraulic actuated piston clutchvehicle transmission. The dual hydraulic actuated piston clutch vehicletransmission includes a gearbox 988 and drive shaft 990. In thisimplementation, the gearbox 988 includes gears 1, 3, and 5 that areassociated with a first clutch of the dual hydraulic actuated pistonclutch vehicle transmission. The gearbox 988 also includes gears 2, 4,and 6 that are associated with a second clutch of the dual hydraulicactuated piston clutch vehicle transmission. The power generated by theengine flows through the transmission and the drive shaft 990 beforereaching the drive wheels 992. The gearbox 988 controls the speed andtorque from the engine that is available to the drive wheels 992.

Referring now to FIG. 16, an exploded view of the radial hydraulicpiston-actuated torque transfer device 900 is shown with variouscomponents that were previously discussed with reference to FIG. 9A.Specifically, FIG. 16 shows the construction/assembly of the pistons 912and rolling engagement elements 914, as well as the hydraulic cylinders910. Additionally, the fixed housing 920 and the seals 924 and 926 ofthe radial hydraulic piston-actuated torque transfer device 900 areshown in this exploded view.

In addition to automotive applications, the present torque transferdevice can be used in various power shaft coupling applicationsinvolving under load actuation of working assemblies including, forexample, construction devices, forestry and agricultural vehicles anddevices, and stationary applications including drilling devices andwinches.

Referring now to FIG. 17, a sectional view of the torque transfer deviceis shown. The torque transfer device with an intermediate control systemis able to transfer torque between two rotating elements, as well ascouple and decouple the rigid connection between them in a controlledmanner. The torque transfer device is also able to dampen torsionaloscillations and improve the smoothness of a coupling process. Thesystem is able to transfer high torques with a short packaging and lowmass, low complexity technical solutions compared to conventionalsolutions. Additionally, the system has reduced manufacturing cost dueto less complexity, and is scalable for larger applications. The systemalso has multiple applications inside a gearbox to engage different gearsets.

As shown in FIG. 17, in order to transfer torque in a controlled mannerfrom an input shaft 1702 to an output shaft 1752, the input shaft 1702sustained by bearing 1704 is coupled by fastening members 1706 to aninput disc 1708 that has attached, e.g., by welding, an input ring 1710.The assembly of the input disc 1708 and input ring 1710 functions as aflywheel and has attached a gear 1712 to drive or be driven, e.g., atstart-up of an internal combustion engine.

An output disc/body 1714 of the torque transfer device includes ahydraulic cylinder 1716 and a slidably engaged hydraulic piston 1718(offset radially) including a rolling engagement element 1720. Thehydraulic cylinder 1716 is connected to a hydraulic circuit via a radialconduit 1730, axial conduit 1732 and radial conduit 1728 within theshaft. The housing conduit 1724 within housing 1726 that is associatedwith hydraulic connector 1722 assures a connection to the hydrauliccircuit.

The torque transfer device with intermediate control system furtherincludes a bearing 1734 having an annular channel 1736 to assure acontinuous connection to radial conduit 1728. The axial conduit 1732 islocked by the cap 1740. The output disc 1714 is sustained within theinput disc 1708 by gliding bearings. A separate bearing 1756 ispositioned between the outer disc assembly 1714 and the housing 1726.The hydraulic circuit comprises hydraulic pipe 1742, control valve 1744,and hydraulic accumulator 1746 for intermediate energy storage and rapidsystem response, having the additional role of damping oscillations ofthe hydraulic fluid. The hydraulic fluid is pumped by hydraulic pump1748 from tank 1750.

The accumulator performs the function of damping and reduction of waveamplitudes, by accumulating (i.e., removing) energy of a pressure peak(relative to a mean pressure level) and releasing the energy in apressure low (relative to a mean pressure level). Waves are generated bysystem actuation. Constant flow is needed to provide accurate control.Control may be disrupted by reflecting and interfering waves.Accordingly, the positioning of an accumulator in the hydraulic controlcircuit enables the accumulator to eliminate perturbation effects in themain control system. Notably, in this implementation, the accumulatorfor the clutch has pressure damping function, not an energy storagefunction.

In this manner, the torque transfer device described herein increasesefficiency as a result of the lower drag torque relative to aconventional system. Notably, the clutch of the system described hereinprovides the technological improvements of: very high power density,improved packaging, lower mass and complexity, and consequently lowercost.

As shown in FIG. 18, the torque transfer device transfers torque in acontrolled manner from an input shaft 1802 to an output shaft 1852. Theinput shaft 1802 is sustained by bearings 1804 and is coupled byfastening members 1806 to an input disc 1808 that is attached, forexample, by welding, to an input ring 1810. The assembly of the inputdisc 1808 and input ring 1810 functions as a flywheel and has attached agear 1812 to drive or be driven, for example, at the start-up of aninternal combustion engine.

An output disc/body 1814 of the torque transfer device includeshydraulic cylinders 1816 and slidably engaged hydraulic pistons 1818,which are associated with rolling engagement element 1820. The hydrauliccylinders 1816 are connected to a hydraulic circuit via a radialconduit, axial conduit, and radial conduit within the shaft. A housingconduit within housing 1826 is associated with the hydraulic connector1822, and assures a connection to the hydraulic circuit. The hydraulicfluid flows along the hydraulic fluid path 1860 from the housing conduitwithin the hydraulic connector 1822, through the dial conduit within theshaft, the axial conduit, the radial conduit, and into the hydrauliccylinders 1816 to actuate the hydraulic pistons 1818 and urge theassociated rolling engagement elements 1820 against the input ring 1810.

The torque transfer device further includes a bearing 1834 having anannular channel 1836 to assure a continuous connection to the radialconduit. The axial conduit is locked by the cap 1840. The output disc1814 is sustained within the input disc 1808 by gliding bearings. Aseparate bearing 1856 is positioned between the outer disc assembly 1814and the housing 1826. When the rolling engagement elements 1820 haveengaged the input ring 1810, the torque flows along the torque path 1870from the input ring 1810, through the input disc 1808, input ring 1810,the rolling engagement elements 1820, the hydraulic pistons 1818, andinto the output disc 1814 and the output shaft 1852.

Referring now to FIG. 19, the torque transfer device transfers torque ina controlled manner from an input shaft to an output shaft. An outputdisc/body 1914 of the torque transfer device includes hydrauliccylinders and slidably engaged hydraulic pistons 1918 (offset radially),which are associated with rolling engagement elements 1920. Thehydraulic cylinders are connected to a hydraulic circuit via a feedingchannel 1916, radial conduit 1930, and axial conduit 1932. Due to theradial offset position of the hydraulic pistons 1918, the hydraulicpistons 1918 are locked, and transfer torque from the input disc 1910 tothe output disc 1914 as long the hydraulic pistons 1918 are extended byhydraulic fluid pressure. In order to balance the torque transferdevice, the hydraulic pistons 1918 are equally spaced in a symmetricmanner about the inner periphery of the input ring 1910, as shown inFIG. 19. To create additional torque, additional pistons 1918 and rollerengagement elements 1920 may be added around the inner periphery of theinput ring 1910, and are actuated simultaneously, without increasing theoverall width of the system. The supplementation of additional pistons1918 and roller engagement elements 1920 also assures ease ofmanufacture scalability.

FIG. 20 shows the application of the present torque transfer device in adual clutch configuration using separate torque transferdevices/clutches for odd and even gear sets, to provide a fast-shifting,smooth operating transmission. This implementation of the torquetransfer device includes a housing 2010, an input disc shaft and ringassembly 2020, a first radial piston torque transfer assembly 2030, afirst hydraulic fluid conduit 2040, a second radial piston torquetransfer assembly 2050, a second hydraulic fluid conduit 2060, a firstoutput disc shaft and ring assembly 2070, and a second output disc shaftand ring assembly 2080. The first radial piston torque transfer assembly2030 contains multiple hydraulic cylinders 2032 and pistons 2034 thatare actuated by pressurized hydraulic fluid. The pistons 2034 each havemounted roller engagement elements 2036 that are urged into contact withthe inner periphery of a ring portion of the input disc shaft and ringassembly 2020. When the roller engagement elements 2036 associated withthe first radial piston torque transfer assembly 2030 are forced intopressurized contact with the input disc shaft and ring assembly 2020,then torque is transferred from the input disc shaft and ring assembly2020 to the first output disc shaft and ring assembly 2070.

Correspondingly, the second radial piston torque transfer assembly 2050contains multiple hydraulic cylinders 2052 and pistons 2054 that areactuated by pressurized hydraulic fluid. The pistons 2054 each havemounted roller engagement elements 2056 that are urged into contact withthe inner periphery of a ring portion of the input disc shaft and ringassembly 2020. When the roller engagement elements 2056 associated withthe second radial piston torque transfer assembly 2050 are forced intopressurized contact with the input disc shaft and ring assembly 2020,then torque is transferred from the input disc shaft and ring assembly2020 to the second output disc shaft and ring assembly 2080.

FIG. 21 is a perspective sectional view of a hydraulic piston torquetransfer device for large torque applications and lower pressure levels.For such high-torque/low-pressure implementations, a larger number ofrows of pistons (e.g., five layers of pistons in FIG. 21) are employedto generate high transmission torque. This implementation of torquetransfer device includes an input shaft 2110, an input disc 2120,actuation gears 2124, hydraulic pistons 2130, rolling engagementelements 2134, hydraulic fluid supply channels 2138, an input ring 2140,bearings 2144, an output cap 2150, a fixed hydraulic fluid supplyconnector 2160, a housing 2170, and an output shaft 2180. The actuationgears 2124 may attach to additional mechanical couplings, such as anengine starter. The torque transfer device contains multiple hydraulicpistons 2130 and rolling engagement elements 2134 that are actuated bypressurized hydraulic fluid that travels through the hydraulic fluidsupply channels 2138 and the fixed hydraulic fluid supply connector2160. The hydraulic pistons 2130 each have mounted roller engagementelements 2134 that are urged into contact with the inner periphery ofthe input ring 2140, which is connected to the input disc 2120 and theinput shaft 2110. When the roller engagement elements 2134 are forcedinto pressurized contact with the input ring 2140, then torque istransferred from the input shaft 2110 to the output shaft 2180.

FIG. 22 is a perspective view of hydraulic piston torque transfer devicefor large torque applications and increased pressure levels. Forhigh-torque/high-pressure implementations, a smaller number of rows ofpistons are needed to generate the necessary transmission torque. Forone such high-torque/high-pressure implementations shown in FIG. 22,three layers of pistons are employed that generate high transmissiontorque. Similarly to the implementation shown in FIG. 21, thisimplementation of the torque transfer device includes an input shaft(not shown), an input disc (not shown), hydraulic pistons 2230, rollingengagement elements 2234, hydraulic fluid supply channels (not shown), aring (not shown), a fixed hydraulic fluid supply connector 2260, ahousing 2270, and an output shaft 2280. The torque transfer devicecontains multiple hydraulic pistons 2230 and rolling engagement elements2234 that are actuated by pressurized hydraulic fluid that travelsthrough the hydraulic fluid supply channels and the fixed hydraulicfluid supply connector 2260. The hydraulic pistons 2230 each havemounted roller engagement elements 2234 that are urged into contact withthe inner periphery of the supporting ring (not shown), which isconnected to the input disc and the input shaft. When the rollerengagement elements 2234 are forced into pressurized contact with theinput ring (not shown), then torque is transferred from the input shaftto the output shaft 2280.

Referring now to FIG. 23, an exploded perspective view of the multiplepiston torque transfer device is shown that relates to theimplementation in FIG. 4C. For implementations in which a hollow spaceis needed inside the torque transfer system, the positioning of thepistons remains the same, while changes are made only to the actuationpath of the hydraulic fluid, in order to be directed from the outsideinstead of the inside. This torque transfer device has a designstructure dedicated to external feeding, as shown in FIG. 23. The torquetransfer device includes an input hub 2310 and an output hub 2320designed to carry the transformer piston assembly 2330. The transformerpiston assembly 2330 includes rolling engagement elements 2334 that areactuated by pressurized hydraulic fluid into contact with the innerperiphery of the input hub 2310. The output hub 2320 also includes anannular feeding channel 2340. The torque transfer device includes afixed hydraulic liquid feeding ring 2350 that contains the annularfeeding channel 2360 and an external feeding hydraulic connection 2370.

The output hub 2320 further includes an orifice 2380 that connects inrotating displacement the hydraulic channels 2340 and 2360. The orifice2380 and the hydraulic channels 2340 and 2360 provide the pathway fromthe hydraulic fluid that is used to actuate the rolling engagementelements 2334 of the transformer piston assembly 2330. The output hub2320 contains a feeding channel 2438 (shown in FIG. 24) that is alignedwith the annular feeding channel 2360 of the fixed hydraulic liquidfeeding ring 2350, so that there is a continuous, permanent flow pathfor the oil, from the annular feeding channel 2360 to the feedingchannel 2438.

FIGS. 24A and 24B illustrate a front view of the multiple piston torquetransfer device shown in FIG. 23, in which a hollow space is neededinside the torque transfer system. In FIG. 24A, the torque transferdevice includes an input hub 2410 and an output hub 2416 that isconfigured to carry the hydraulic cylinders 2420 and the transformerpistons 2430. The transformer pistons 2430 include rolling engagementelements 2434 that are actuated by pressurized hydraulic fluid intocontact with the inner periphery of the input hub 2410. The output hub2416 also includes an annular feeding channel 2440. Referring now toFIG. 24B, the torque transfer device includes a fixed hydraulic liquidfeeding ring 2450 that contains the annular feeding channel 2460 and anexternal feeding hydraulic connection 2470. As described above withrespect to FIG. 23, the annular feeding channel 2440 is connected to thefeeding space 2438 that assures the hydraulic path actuates thetransformer pistons 2430 by flowing through annular feeding channel2440.

In any of the clutches and torque transfer devices described herein, thetorque transferred by the clutch or the torque transfer device isdirectly proportional to a pressure exerted against a piston or aplurality of pistons of the clutch or of the torque transfer device.Targets for many clutches and torque transfer devices include thecapability to transfer up to 150 Nm of torque, to rotate at up to 5500rpm, to operate at 300-600 psi, and to have a coupling time between 0.1and 0.5 seconds. Prior clutches and torque transfer devices have beencapable of transferring up to 160 Nm of torque, rotating at up to 6000rpm, operating at up to 580 psi, and having a coupling time between 0.1and 0.5 seconds. The clutches and torque transfer devices describedherein are capable of transferring up to 1500 Nm of torque, rotating atup to 10000 rpm, operating at up to 2200 psi, and having a coupling timebetween 0.1 and 0.5 seconds. In particular, it has been calculated that,by appropriate selection of a volume of a working fluid, a configurationof a pipe orifice, and a valve actuation time, the clutches and torquetransfer devices described herein are capable of coupling from 0% to100% in 0.1 seconds, although this time can be increased as needed toaccommodate drivability demands.

FIG. 25 illustrates various components of an automobile 2500. Inparticular, FIG. 25 illustrates a pair of wheels 2502 connected by arear axle 2504, a rear differential 2506, an exhaust pipe 2508 includinga waste heat recovery system, and a brake energy recovery system 2510.FIG. 26 illustrates some of the components of the automobile 2500illustrated in FIG. 25, including the exhaust pipe 2508, which has awaste heat recovery system, and the brake energy recovery system 2510,which has a dual high-pressure and low-pressure accumulator 2512.

FIG. 27 illustrates a rear-wheel drive system 2700 of an automobile. Thedrive system 2700 includes an internal combustion engine 2702, a vehiclegearbox 2704 mechanically coupled to an output of the engine 2702, and afirst clutch or torque transfer device 2706 mechanically coupled to anoutput of the gearbox 2704. The drive system 2700 also includes a firsthydraulic pump 2708 (which can be inverted to operate as a hydraulicmotor) mechanically coupled to an output of the clutch 2706, and a firstflow control valve 2710 hydraulically coupled at first and second portsto an inlet and an outlet of the first hydraulic pump 2708.

The drive system 2700 also includes an accumulator 2712 hydraulicallycoupled to third and fourth ports of the first flow control valve 2710,a second flow control valve 2714 hydraulically coupled at first andsecond ports to the accumulator 2712, and a second hydraulic pump 2716(which can be inverted to operate as a hydraulic motor) hydraulicallycoupled to third and fourth ports of the second flow control valve 2714.The drive system 2700 also includes a second clutch or torque transferdevice 2718 mechanically coupled to the second hydraulic pump 2716, adifferential 2720 mechanically coupled to the clutch 2718, an axle 2722mechanically coupled to the differential 2720, and a pair of wheels 2724mounted on the axle 2722.

The accumulator 2712 is a dual-sided accumulator coupled at its firstside to the first flow control valve 2710 and at its second side to thesecond flow control valve 2714. The accumulator 2712 includes paralleland integrated high-pressure and low-pressure accumulators 2713 a and2713 b, respectively. The high-pressure accumulator 2713 a includes acompressible gas 2713 c positioned and sealed between two movablepistons 2713 d.

The first flow control valve 2710 is coupled at its first port to anoutput of the first hydraulic pump 2708, at its second port to an inputof the first hydraulic pump 2708, at its third port to the high-pressureaccumulator 2713 a, and at its fourth port to the low-pressureaccumulator 2713 b. In a first position of the first flow control valve2710, illustrated in FIG. 27, the first and third ports of the firstflow control valve 2710 are coupled to one another to hydraulicallycouple the output of the hydraulic pump 2708 to the high-pressureaccumulator 2713 a. Further, in the first position, the second andfourth ports of the first flow control valve 2710 are coupled to oneanother to hydraulically couple the input of the hydraulic pump 2708 tothe low-pressure accumulator 2713 b. The first flow control valve 2710can be switched from its first position to a second position, in whichthe first and fourth ports couple the output of the hydraulic pump 2708to the low-pressure accumulator 2713 b, and the second and third portscouple the input of the hydraulic pump 2708 to the high-pressureaccumulator 2713 b.

The second flow control valve 2714 is coupled at its first port to thehigh-pressure accumulator 2713 a, at its second port to the low-pressureaccumulator 2713 b, at its third port to an output of the secondhydraulic pump 2716, and at its fourth port to an input of the secondhydraulic pump 2716. In a first position of the second flow controlvalve 2714, illustrated in FIG. 27, the first and third ports of thesecond flow control valve 2714 are coupled to one another tohydraulically couple the output of the hydraulic pump 2716 to thehigh-pressure accumulator 2713 a. Further, in the first position, thesecond and fourth ports of the second flow control valve 2714 arecoupled to one another to hydraulically couple the input of thehydraulic pump 2716 to the low-pressure accumulator 2713 b. The secondflow control valve 2714 can be switched from its first position to asecond position, in which the first and fourth ports couple the input ofthe hydraulic pump 2716 to the high-pressure accumulator 2713 a and thesecond and third ports couple the output of the hydraulic pump 2716 tothe low-pressure accumulator 2713 b.

The first and second hydraulic pumps 2708 and 2716 have any suitableconfiguration known in the field of hydraulic pumps that allows them tobe reversed or inverted to act as hydraulic motors, with vane-typehydraulic pumps being one example. The first and second clutches 2706and 2718 have configurations matching any of the clutches or torquetransfer devices described herein. The gearbox 2704 has a configurationmatching any of the gearboxes described herein. The engine 2702, thedifferential 2720, axle 2722, and wheels 2724 have well-known andconventional configurations.

The drive system 2700 replaces a traditional drive shaft coupling agearbox to a differential. Thus, during operation of the drive system2700, power generated by the engine 2702 can be transferred through thegearbox 2704, the first clutch 2706, the first hydraulic pump 2708, thefirst flow control valve 2710, the accumulator 2712, the second flowcontrol valve 2714, the second hydraulic pump 2716, the second clutch2718, the differential 2720, the axle 2722, and to the wheels 2724. Theengine 2702 can be run continuously at its most efficient operatingparameters. When power demanded at the wheels 2724 is greater than powersupplied by the engine 2702, the drive system 2700 releases some of theenergy stored in the accumulator 2712. When power demanded at the wheels2724 is less than power supplied by the engine 2702, the drive system2700 stores some of the energy provided by the engine 2702 in theaccumulator 2712. The drive system 2700 provides a continuously variabletransmission capable of providing any required power to the wheels 2724independent of the power supplied by the engine 2702.

During deceleration, the valves 2710 and 2714 are turned to theirrespective first positions and the wheels 2724 are used as a source ofmechanical power that drives operation of the first and second hydraulicpumps 2708 and 2716 to pump fluid from the low-pressure accumulator 2713b to the high-pressure accumulator 2713 a, to recover and store energyfor later. The drive system 2700 thus recovers energy from the wheels2724 and stores it in the accumulator 2712. During acceleration, thevalves 2710 and 2714 are turned to their respective second positions andthe pumps 2708 and 2716 are inverted to operate as hydraulic motors.Thus, high-pressure fluid held within the high-pressure accumulator 2713a is released to flow through the motors to the low-pressure accumulator2713 b to drive operation of the hydraulic motors to increase therotational speed of the wheels 2724. The drive system 2700 thusdischarges energy from the accumulator 2712 to drive the wheels 2724. Insuch an implementation, the accumulator 2712 recovers energy from, orprovides energy to, the wheels 2724 depending on the circumstances,improving overall efficiency.

FIG. 28 illustrates a front-wheel drive system 2800 of an automobile.The drive system 2800 includes an internal combustion engine 2802, avehicle gearbox 2804 mechanically coupled to an output of the engine2802, a front axle 2806 mechanically coupled to an output of the gearbox2804, and a pair of wheels 2808 mounted on the axle 2806. The drivesystem 2800 also includes an additional drive gear 2810 mechanicallycoupled to the front axle 2806, a clutch or torque transfer device 2812mechanically coupled to the additional drive gear 2810, a hydraulic pump2814, which can be inverted to operate as a hydraulic motor,mechanically coupled to an output of the clutch 2812, and a flow controlvalve 2816 hydraulically coupled at first and second ports to an inletand an outlet of the hydraulic pump 2814.

The drive system 2800 also includes an accumulator 2818 hydraulicallycoupled to third and fourth ports of the first flow control valve 2816.The accumulator 2818 is a single-sided accumulator coupled at its singleside to the flow control valve 2816. The accumulator 2818 includesparallel and integrated high-pressure and low-pressure accumulators 2820a and 2820 b, respectively. The drive system 2800 functions in the samemanner as the drive system 2700 except that it is a front-wheel drivesystem, it includes a single-sided accumulator 2818 coupled to the frontaxle 2806 by the additional drive gear 2810, and it does not replace anymechanical connections between any of the other components of theautomobile (e.g., it does not replace a traditional drive shaft).

FIG. 29 illustrates a front-wheel drive system 2900 of an automobile.The drive system 2900 includes an internal combustion engine 2902, avehicle gearbox 2904 mechanically coupled to an output of the engine2902, a front axle 2906 mechanically coupled to an output of the gearbox2904, and a pair of wheels 2908 mounted on the axle 2906. The drivesystem 2900 also includes an extension 2910 of a shaft of the gearbox2904, a clutch or torque transfer device 2912 mechanically coupled tothe extension 2910 of the shaft of the gearbox 2904, a hydraulic pump2914, which can be inverted to operate as a hydraulic motor,mechanically coupled to an output of the clutch 2912, and a flow controlvalve 2916 hydraulically coupled at first and second ports to an inletand an outlet of the hydraulic pump 2914.

The drive system 2900 also includes an accumulator 2918 hydraulicallycoupled to third and fourth ports of the first flow control valve 2916.The accumulator 2918 is a single-sided accumulator coupled at its singleside to the flow control valve 2916. The accumulator 2918 includesparallel and integrated high-pressure and low-pressure accumulators 2920a and 2920 b, respectively. The drive system 2900 functions in the samemanner as the drive system 2700 except that it is a front-wheel drivesystem, it includes a single-sided accumulator 2918 coupled to thegearbox 2904 by the extension 2910 of the shaft of the gearbox 2904, andit does not replace any mechanical connections between any of the othercomponents of the automobile (e.g., it does not replace a traditionaldrive shaft).

FIG. 30 illustrates a front-wheel drive system 3000 of an automobile.The drive system 3000 includes an internal combustion engine 3002, avehicle gearbox 3004 mechanically coupled to an output of the engine3002, a front axle 3006 mechanically coupled to an output of the gearbox3004, and a pair of wheels 3008 mounted on the axle 3006. The drivesystem 3000 also includes an additional drive gear 3010 mechanicallycoupled to the front axle 3006, a clutch or torque transfer device 3012mechanically coupled to the additional drive gear 3010, a hydraulic pump3014, which can be inverted to operate as a hydraulic motor,mechanically coupled to an output of the clutch 3012, and a flow controlvalve 3016 hydraulically coupled at first and second ports to an inletand an outlet of the hydraulic pump 3014.

The drive system 3000 also includes an accumulator 3018 hydraulicallycoupled to third and fourth ports of the first flow control valve 3016.The accumulator 3018 is a dual-sided accumulator coupled at its firstside directly to the flow control valve 3016 and at its second side tothe flow control valve 3016 via a second flow control valve 3020. Theaccumulator 3018 includes parallel and integrated high-pressure andlow-pressure accumulators 3022 a and 3022 b, respectively. The drivesystem 3000 functions in the same manner as the drive system 2800 exceptthat it includes a dual-sided accumulator 3018 and a second flow controlvalve 3020 rather than a single-sided flow control valve.

The drive system 3000 includes the additional drive gear 3010 and theclutch 3012 coupled to the additional drive gear 3010, in a mannersimilar to that described above for drive system 2800. In an alternativeimplementation, the drive system 3000 includes an extension of a shaftof the gearbox 3004, and the clutch 3012 is mechanically coupled to theextension, in a manner similar to that described above for drive system2900.

FIG. 31A illustrates a conventional (e.g., prior art) rear-wheel drivesystem 3100 of an automobile. The drive system 3100 includes a gearbox3102, a rear differential 3104, and a drive shaft 3106 mechanicallycoupling the gearbox 3102 to the rear differential 3104. FIG. 31Billustrates the rear-wheel drive system 3100 of FIG. 31A with a brakeenergy recovery system 3108 in place of the drive shaft 3106. The brakeenergy recovery system 3108 has features matching those of the drivesystem 2700 described above with respect to FIG. 27.

FIG. 32A illustrates a conventional (e.g., prior art) all-wheel drivesystem 3200 of an automobile. The drive system 3200 includes a gearbox3202, a front differential 3204, a front drive shaft 3206 mechanicallycoupling the gearbox 3202 to the front differential 3204, a reardifferential 3208, and a rear drive shaft 3210 mechanically coupling thegearbox 3202 to the rear differential 3208. FIG. 32B illustrates therear-wheel drive system 3200 of FIG. 32A with a brake energy recoverysystem 3212 in place of the drive shaft 3210. The brake energy recoverysystem 3212 has features matching those of the drive system 2700described above with respect to FIG. 27.

FIG. 32C illustrates a rear wheel drive system coupled to a DoubleActing Brake Energy Recovery System with a structure having anintegrated direct mechanical transmission. In FIG. 32D, for a DoubleActing Brake Energy Recovery System, the structure includes anintegrated direct mechanical transmission for vehicles that operate athigh speeds (velocity) and for vehicles in which efficiency reductioninduced by hydraulic losses must be avoided (e.g., school buses, parcelservice vehicles, delivery vans and pick-ups, taxis, and the like). Thehydrostatic run mode enables low load/low speed operating conditions tobe achieved at optimum engine map points (curves) by continuouslystoring and releasing energy.

Referring now to FIG. 32D, a rear wheel drive system coupled with aDouble Acting Brake Energy Recovery System is illustrated with astructure having an integrated power split transmission. In FIG. 32D,for a Double Acting Brake Energy Recovery System, the structure includesan integrated power split transmission for vehicles that require largerranges of mechanical gear ratios coupled with continuous transmissionratios control, which allows for optimum engine map run (e.g., heavyduty vehicles, construction, forestry vehicles, and the like). Thestructure in FIG. 32D relies on conventional gear sets.

For all wheel drive applications, the integration of Brake EnergyRecovery System relies on replacing the propulsion shaft of theall-wheel drive vehicle propulsion systems that drive the rear axle,with a double acting brake energy recovery system in the followingdifferent structural versions. FIG. 32F illustrates an all-wheel drivesystem coupled Double Acting Brake Energy Recovery System. FIG. 32Gillustrates an all-wheel drive system coupled Double Acting Brake EnergyRecovery System with a structure having an integrated hydrostatictransmission. FIG. 32H illustrates an all-wheel drive system coupledDouble Acting Brake Energy Recovery System with a structure having anintegrated direct mechanical transmission. FIG. 32I illustrates anall-wheel drive Drive system coupled Double Acting Brake Energy RecoverySystem with a structure having an integrated power split transmissionand conventional gear sets. FIG. 32J illustrates an all-wheel driveDrive system coupled to Double Acting Brake Energy Recovery System witha structure having an integrated power split transmission and planetarygear sets.

FIG. 33 illustrates a variety of operating conditions for a brake energyrecovery system 3300 coupled to a drive shaft 3302 of a drive system ofan automobile. In particular, FIG. 33 illustrates that as the automobileis driving at a constant speed, such as while cruising on a freeway,power generated by an internal combustion engine is transmittedmechanically by the drive shaft 3302 to wheels of the vehicle, as shownat 3304 at the top of FIG. 33. As the automobile brakes, the drive shaft3302 is decoupled, such as at a clutch, and the wheels of the vehicleare used to power operation of hydraulic pumps of the brake energyrecovery system 3300 to store energy in an accumulator of the brakeenergy storage system 3300, such as in the manner described above withrespect to FIG. 27, as shown at 3306.

As the automobile accelerates, the drive shaft 3302 is decoupled, suchas at a clutch, and the energy stored in the accumulator of the brakeenergy storage system 3300 is released to power hydraulic motors of thebrake energy storage system 3300 to power the wheels of the vehicle,such as in the manner described above with respect to FIG. 27, as shownat 3308. It has been found that such storage and reuse of stored energyis about 81% efficient. As the automobile is driving in stop-and-goconditions, such as while driving through a city, an internal combustionengine is operated under its most efficient operating parameters togenerate a constant amount of power, while the drive shaft 3302 isdecoupled, such as at a clutch, and the brake energy recovery system3300 is used as needed to store excess produced energy or to releaseenergy to meet excess power demand, as shown at 3310 at the bottom ofFIG. 33. It has been found that such applications improve overall systemefficiency by 5-10 percent depending on the running conditions.

FIG. 34A illustrates a rear-wheel drive system 3400 of a heavy-dutyautomobile or vehicle, such as a military vehicle, police vehicle, firevehicle such as a fire truck or fire engine, construction vehicle,forestry vehicle, agricultural vehicle such as an agricultural tractor,semi-truck, delivery van, parcel service vehicle, off-road vehicle,school bus, forklift, pickup truck, taxi, and the like. The drive system3400 includes a first drive shaft 3402 mechanically coupled at a firstend thereof to an output of a gearbox 3404 and a second drive shaft 3406mechanically coupled at a first end thereof to a second end of the firstdrive shaft 3402 and at a second end thereof to a differential 3408.Thus, the first and second drive shafts 3402 and 3406 are mechanicallycoupled to one another in series, and together couple the gearbox 3404to the differential 3408, but can be decoupled as needed, such asdescribed above with respect to FIG. 33.

FIG. 34B illustrates two brake energy recovery systems 3410 coupled tothe first drive shaft 3402 of the drive system 3400. The brake energyrecovery systems 3410 have the same features, and operate in the samemanner, as the brake energy recovery system 2700 and/or the brake energyrecovery system 3300 described above. Respective first ends of the brakeenergy recovery systems 3410 are coupled to the first drive shaft 3402by a first set of drive gears 3412, and respective second ends of thebrake energy recovery systems 3410 are coupled to the first drive shaft3402 by a second set of drive gears 3414. As also illustrated in FIG.34B, the drive system 3400 has a cross-sectional shape including a firstelliptical housing 3416 for a first one of the brake energy recoverysystems 3410, a second elliptical housing 3418 for a second one of thebrake energy recovery systems 3410, and a circular housing 3420 for thefirst drive shaft 3402. Providing the drive system 3400 with the brakeenergy recovery systems 3410 improves its effectiveness for use inheavy-duty vehicles.

FIG. 34C illustrates three brake energy recovery systems 3410 coupled tothe first drive shaft 3402 and three brake energy recovery systems 3410coupled to the second drive shaft 3406 of the drive system 3400. Thebrake energy recovery systems 3410 have the same features, and operatein the same manner, as the brake energy recovery system 2700 and/or thebrake energy recovery system 3300 described above. Respective first endsof the three brake energy recovery systems 3410 coupled to the firstdrive shaft 3402 are coupled to the first drive shaft 3402 by a firstset of drive gears 3412, and respective second ends of the brake energyrecovery systems 3410 coupled to the first drive shaft 3402 are coupledto the first drive shaft 3402 by a second set of drive gears 3414.Respective first ends of the three brake energy recovery systems 3410coupled to the second drive shaft 3406 are coupled to the second driveshaft 3406 by a third set of drive gears 3422, and respective secondends of the brake energy recovery systems 3410 coupled to the seconddrive shaft 3406 are coupled to the second drive shaft 3406 by a fourthset of drive gears 3424.

As illustrated in FIG. 34C, the drive system 3400 has twocross-sectional shapes including an elliptical housing 3426 that housesfirst, second, and third circular housings for respective first, second,and third brake energy recovery systems 3410. Providing the drive system3400 with the brake energy recovery systems 3410 improves itseffectiveness for use in heavy-duty vehicles.

FIG. 35A illustrates a conventional (e.g., prior art) all-wheel drivesystem 3500 of a heavy-duty automobile or vehicle, such as any of theheavy-duty vehicles described above with respect to rear-wheel drivesystem 3400. The drive system 3500 includes a gearbox 3502, a frontdrive shaft 3504 mechanically coupled at a first end thereof to anoutput of the gearbox 3502, a front differential 3506 mechanicallycoupled to a second end of the front drive shaft 3504, and a front axle3508 mechanically coupled to the front differential 3506. The drivesystem 3500 also includes a rear drive shaft 3510 mechanically coupledat a first end thereof to an output of the gearbox 3502, a reardifferential 3512 mechanically coupled to a second end of the rear driveshaft 3510, and a rear axle 3514 mechanically coupled to the reardifferential 3512.

FIG. 35B illustrates a front brake energy recovery system 3516 coupledto the front axle 3508 and a rear brake energy recovery system 3518coupled to the rear axle 3514. The front and rear brake energy recoverysystems 3516 and 3518 have the same features, and operate in the samemanner, as the brake energy recovery system 2700, drive system 2800,drive system 2900, and/or drive system 3000 described above. First andsecond ends of the front brake energy recovery system 3516 are coupledto the front axle 3508 by front gear sets 3520 and first and second endsof the rear brake energy recovery system 3518 are coupled to the rearaxle 3514 by rear gear sets 3522. The all-wheel drive system 3500 issimilar in many respects to the all-wheel drive system 3200, butprovides two brake energy recovery systems each coupled to a respectiveaxle, rather than one brake energy recovery system in place of a driveshaft.

FIG. 36 illustrates electrical connections of a brake energy recoverysystem 3600 in place of an axle between two wheels 3602 of anautomobile. The brake energy recovery system 3600 has the same features,and operate in the same manner, as the brake energy recovery system 2700described above, except that it is coupled directly to wheels of theautomobile, rather than between a gearbox and a differential thereof.Brake energy recovery system 3600 includes a first hydraulic pump 3604,a first flow control valve 3606, an accumulator 3608 including ahigh-pressure accumulator 3610 having a pressurized gas 3612, a secondflow control valve 3614, and a second hydraulic pump 3616. The brakeenergy recovery system 3600 also includes a battery 3618 that supplieselectrical power to a first actuator 3620 for the first hydraulic pump3604, a second actuator 3622 for the first flow control valve 3606, athird actuator 3624 for the second flow control valve 3614, and a fourthactuator 3626 for the second hydraulic pump 3616.

The brake energy recovery system 3600 also includes a control unit 3628that is electrically and communicatively coupled to receive input froman accelerator or gas pedal 3630 of the automobile, a brake pedal 3632of the automobile, and a pressure transducer 3634 measuring a pressureof the pressurized gas 3612. The control unit 3628 is also electricallyand communicatively coupled to provide commands to the first, second,third, and fourth actuators 3620, 3622, 3624, and 3626.

FIG. 37A illustrates a brake energy recovery system 3700 in place of anaxle between two wheels 3702 of an automobile. The brake energy recoverysystem 3700 has the same features, and operate in the same manner, asthe brake energy recovery system 3600 described above. The brake energyrecovery system 3700 includes a thermal energy recovery subsystem 3704that includes an annular fluid jacket 3706 that extends around anexhaust pipe 3708 of the vehicle. The thermal energy recovery subsystem3704 also includes a heat exchanger and is capable of extracting andrecovering heat energy from the exhaust carried through the exhaust pipe3708, to power and/or control operation of the brake energy recoverysystem 3700. FIG. 37B illustrates the brake energy recovery system 3700incorporated into an all-wheel drive system. FIG. 37C illustrates thebrake energy recovery system 3700 incorporated into a rear-wheel drivesystem.

FIG. 38 illustrates a flow diagram of operation of any one of the brakeenergy recovery systems described herein. FIG. 39 illustrates a blockdiagram of operation of a brake energy recovery system, duringacceleration, braking, or constant-speed cruising of a vehicle. Inparticular, after receiving a set of inputs and determining anappropriate course of action, a control unit transmits commands orinstructions to flow control valves of one or more brake energy recoverysystems to turn the valves to positions that allow accumulator(s) topower acceleration of the vehicle, the wheels to power storage of energyin accumulator(s), or neither. Thus, the control unit controls one ormore brake energy recovery systems to operate in the manner describedabove for any of the brake energy recovery systems described herein.

FIG. 40 illustrates a block diagram of a control unit 4000 for a brakeenergy recovery system. In particular, the control unit 4000 isconfigured to accept inputs from a brake pedal at 4002, an acceleratoror gas pedal at 4004, a pressure transducer measuring a pressure withina hydraulic accumulator at 4006, an anti-lock braking system and/or anelectronic stability control system at 4008, data regarding comfortableand safe driving conditions and parameters at 4010, a travelling speedof the vehicle at 4012, a power demanded at 4014, and information fromfirst and second brake energy recovery systems at 4016. The control unit4000 is also configured to transmit commands to a conventional brakingsystem at 4018 and to first and second brake energy recovery systems at4020.

FIG. 41 illustrates another block diagram of operation of a brake energyrecovery system, during acceleration, braking, or constant-speedcruising of a vehicle. In particular, after receiving a set of inputsincluding a pressure of a high-pressure accumulator, the position of gasand/or brake pedals of the vehicle, and information from anti-lock brakeor other systems of the automobile, a control unit determines anappropriate course of action and then transmits commands or instructionsto flow control valves and/or hydraulic pumps/motors of one or morebrake energy recovery systems so that accumulator(s) power accelerationof the vehicle, wheels of the vehicle power storage of energy inaccumulator(s), or neither. Thus, the control unit controls one or morebrake energy recovery systems to operate in the manner described abovefor any of the brake energy recovery systems described herein.

FIG. 42 illustrates the results of an analysis of efficiencyimprovements provided by a brake energy recovery system such as thebrake energy recovery system 2700 described above. In particular, theanalysis shows that the brake energy recovery systems described hereinimprove efficiency of a vehicle in city driving conditions by about 35%.

The following related applications are hereby incorporated herein byreference in their entireties: (1) Hydraulic Actuated Piston Clutch,U.S. Ser. No. 62/498,349, filed Dec. 21, 2016; (2) Radial OffsetHydraulic Piston Torque Transfer System, U.S. Ser. No. 62/605,291, filedAug. 7, 2017; (3) Hybrid Kinematic Hydraulic Transmission for Use withan Integrated Brake Energy Recovery System, U.S. Ser. No. 62/605,283,filed Aug. 7, 2017; (4) Brake Energy Active Recovery System forVehicles, U.S. Ser. No. 62/606,522, filed Sep. 26, 2017; (5) Gear Boxwith Integrated Brake Energy Recovery System, U.S. Ser. No. 62/584,650,filed Nov. 10, 2017; (6) U.S. Ser. No. 62/598,364, filed Dec. 13, 2017;(7) Offset Radial Piston Torque Transfer Device, U.S. Ser. No.62/598,366, filed Dec. 13, 2017; (8) Integrated Brake and Thermal EnergyRecovery System, U.S. Ser. No. 15/731,267, filed May 15, 2017; (9)Radial Hydraulic Piston Actuated Torque Transfer Device, U.S. Ser. No.15/731,271, filed May 15, 2017; and (10) Axial Piston VariableDisplacement Hydraulic Rotational Unit with Integrated Propulsion Shaft,U.S. Ser. No. 15/731,383, filed Jun. 5, 2017.

Additionally, U.S. provisional patent application Nos. 62/496,784, filedOct. 28, 2016; 62/498,348, filed Dec. 21, 2016; 62/498,347, filed Dec.21, 2016; 62/498,338, filed Dec. 21, 2016; 62/498,337, filed Dec. 21,2016; 62/498,336, filed Dec. 21, 2016; 62/606,521, filed Sep. 26, 2017;62/606,511, filed Sep. 26, 2017; 62/577,630, filed Oct. 26, 2017; and62/580,360, filed Nov. 1, 2017; as well as U.S. non-provisional patentapplication Ser. No. 15/731,360, filed Jun. 1, 2017; and PCT applicationno. PCT/US17/58883, filed Oct. 27, 2017, are hereby incorporated hereinby reference in their entireties.

The various embodiments described above can be combined to providefurther embodiments. These and other changes can be made to theembodiments in light of the above-detailed description. In general, inthe following claims, the terms used should not be construed to limitthe claims to the specific embodiments disclosed in the specificationand the claims, but should be construed to include all possibleembodiments along with the full scope of equivalents to which suchclaims are entitled. Accordingly, the claims are not limited by thedisclosure.

The invention claimed is:
 1. A hydraulic transformer clutch employinghydraulic piston assemblies with integrated electrohydraulic actuation,the hydraulic transformer clutch comprising: an output shaft; an outputdisc affixed to the output shaft for rotation therewith; an input shaft;a rotatable housing affixed to one of the input shaft or the outputshaft for rotation therewith; a plurality of hydraulic cylindersoperatively connected to the rotatable housing, the hydraulic cylindersspaced about the rotatable housing, wherein each hydraulic cylinder ofthe plurality of hydraulic cylinders is positioned at a substantiallyconstant offset angle relative to radial directions of the input shaftand the output shaft; and a plurality of working pistons, each piston ofthe plurality of working pistons being slidably mounted within acorresponding hydraulic cylinder of the plurality of hydrauliccylinders, each piston of the plurality of working pistons positioned tobe selectively pushed, when actuated, to create a rigid connectionbetween the input shaft and the output shaft; wherein one or moreactuator pistons are pushed by an electromagnet and create pressure thatis distributed on working piston surfaces and generates active torque.2. The hydraulic transformer clutch of claim 1, further comprising aplurality of rotating engagement elements, each rotating engagementelement of the plurality of rotating engagement elements associated witha piston of the plurality of working pistons, wherein the plurality ofrotating engagement elements engage the output disc when actuated. 3.The hydraulic transformer clutch of claim 1, further comprising ahydraulic system operatively associated with the hydraulic cylinders andworking pistons, wherein the hydraulic system enables actuation andde-actuation of the working pistons in the hydraulic cylinders, whereinactuation of the working pistons in the hydraulic cylinders couples theinput shaft to the output shaft and de-actuation of the working pistonsin the hydraulic cylinders decouples the input shaft from the outputshaft.
 4. The hydraulic transformer clutch of claim 1, furthercomprising a hydraulic variable displacement pump that is operativelyassociated with each hydraulic cylinder and piston.
 5. The hydraulictransformer clutch of claim 4, further comprising directional controlvalves that use hydraulic fluid to selectively urge each piston to bepushed towards the output disc when actuated.
 6. The hydraulictransformer clutch of claim 5, further comprising a hydraulicaccumulator that is operatively associated with the hydraulic variabledisplacement pump and the directional control valves, wherein thehydraulic accumulator reduces oscillations in the hydraulic transformerclutch during actuation.
 7. The hydraulic transformer clutch of claim 1,further comprising a pressure relief valve that protects againstpressure overloads.
 8. The hydraulic transformer clutch of claim 1,wherein the hydraulic transformer clutch is incorporated into anautomotive transmission.